1. Field of the Invention
The present invention relates to improvement of a continuously variable transmission apparatus, for use in an automatic transmission for a vehicle (automobile), in which a toroidal continuously variable transmission is incorporated. Particularly, the invention relates to a continuously variable transmission apparatus to which is added a function of improving characteristics during a vehicle stop or during very low speed driving while implementing a structure which avoids providing a driver with an uncomfortable feeling at vehicle start.
2. Background Art
A toroidal-type continuously variable transmission as shown in FIGS. 7 to 9 has been investigated as an automatic transmission for a vehicle, and has found limited use. The toroidal continuously variable transmission is referred to as a double cavity type, in which input discs 2, 2 are supported on peripheries of both end portions of an input shaft 1 via ball splines 3, 3. Accordingly, the two input discs 2, 2 are rotatably supported such that they are concentric with each other and rotate synchronously. Further, an output gear 4 is supported rotatably with respect to the input shaft 1 on the periphery of the intermediate portion of the input shaft 1. Output discs 5, 5 are splined at respective end portions of a cylindrical portion provided at the center portion of the output gear 4. Hence, the two output discs 5, 5 rotate synchronously with the output gear 4.
A plurality of power rollers 6, 6 (usually two or three power rollers on each side) are interposed between the input discs 2, 2 and the output discs 5, 5. The power rollers 6, 6 are rotatably supported on the inner surfaces of trunnions 7, 7 respectively via support shafts 8, 8 and a plurality of roller bearings. The trunnions 7, 7 are provided so as to swing around pivot shafts 9, 9 disposed for the respective trunnions 7, 7 on both end portions thereof, in longitudinal directions (i.e., in vertical directions in FIGS. 7 and 9, and in a direction perpendicular to the plane of FIG. 8). The trunnions 7, 7 are inclined by hydraulic actuators 10, 10; specifically, the hydraulic actuators 10, 10 displace the trunnions 7, 7 along the axes of pivot shafts 9, 9. Inclination angles of the trunnions 7, 7 are synchronized hydraulically and mechanically.
That is, in the case where the inclination angles of the trunnions 7, 7 are changed in order to change a transmission ratio between the input shaft 1 and the output gear 4, the trunnions 7, 7 are displaced in opposite directions by the actuators 10, 10, respectively. For example, the power roller 6 on the right-hand side in FIG. 9 is displaced downward in FIG. 9, while the power roller 6 on the left-hand side in FIG. 9 is displaced upward in FIG. 9 by the same distance. As a result, forces acting along a tangential direction of the contact portions between the peripheral surfaces of the power rollers 6, 6 and the inner surfaces of the input side discs 2, 2 and the output side discs 5, 5 are changed in direction (in other words, sideslip occurs at contact portions thereof). Consequently, due to the change in direction of the forces, the trunnions 7, 7 swing (incline) in opposite directions around the pivot shafts 9, 9, which are pivotally supported by support plates 11, 11. As a result, contacting portions between the peripheral surfaces of the power rollers 6, 6 and the inner surfaces of the input discs 2, 2 and the output discs 5, 5 are changed. Thereby, a rotation transmission ratio between the input shaft 1 and the output gear 4 changes.
Pressurized oil is supplied to and discharged from the actuators 10, 10 by means of a single control valve 12, irrespective of the number of the actuators 10, 10. The movement of any one of the trunnions 7 is fed back to the control valve 12. The control valve 12 has a sleeve 14 to be displaced in an axial direction (i.e., in the horizontal direction in FIG. 9, and in a direction perpendicular to the plane of FIG. 7) by a stepping motor 13 and a spool 15 fitted into the inner periphery of the sleeve 14 so as to allow displacement in the axial direction thereof. A feedback mechanism is constituted as follows: rods 17, 17 connect the trunnions 7, 7 and pistons 16, 16 of the respective actuators 10, 10; and a precess cam 18 is fixed on an end portion of each of the rods 17 attached to any one of the trunnions 7, 7. The movement of the rod 17; that is, a resultant total of the displacement in the axial direction and the displacement in the rotating direction, is transmitted to the spool 15 via the precess cam 18 and a link arm 19 to thereby displace the spool 15 in the axial direction. A synchronous cable 20 is suspended between the trunnions 7, 7 in such amanner that the inclination angles of the respective trunnions 7, 7 are mechanically synchronized to each other even in the case where trouble arises in a hydraulic system.
At the time of switching the transmission state, the sleeve 14 is displaced to a position corresponding to a desired transmission ratio by the stepping motor 13 to thereby open a flow path to a predetermined direction of the control valve 12. As a result, the pressurized oil is supplied to the actuators 10, 10 in a predetermined direction, whereby the actuators 10, 10 displace the trunnions 7, 7 in a predetermined direction. That is, in accordance with supply of the pressurized oil, the trunnions 7, 7 swing around the pivot shafts 9, 9 while being displaced in the axial direction of the pivot shafts 9, 9. Then, the motion (i.e., the motion in an axial direction and the swing) of one of the trunnions 7 is transmitted to the spool 15 via the precess cam 18 fixed to the end portion of the rod 17 and the link arm 19 to thereby displace the spool 15 in the axial direction. As a result, the flow path of the control valve 12 is closed in a state where the trunnions 7 are displaced by the predetermined amount, and supply and discharge of the pressurized oil to and from the actuators 10, 10 is stopped.
The operation of the control valve 12 based on the displacement of the trunnion 7 and the cam surface 21 of the precess cam 18 during the above is as follows. First, the trunnion 7 is displaced in the axial direction along with the opening of the flow path of the control valve 12. Then, as described hitherto, in response to the sideslip generated on the contact portions between the peripheral surface of the power roller 6 and the inner peripheral surfaces of the input disc 2 and the output disc 5, the trunnion 7 starts swinging around the pivot shaft 9. Further, along with the displacement of the trunnion 7 in the axial direction, a displacement of a cam surface 21 is transmitted to the spool 15 via the link arm 19. Thereby, the spool 15 is displaced in the axial direction and a state of the control valve 12 is changed. More specifically, the control valve 12 is switched, by the actuator 10, in a direction of returning the trunnion 7 to a neutral position.
Accordingly, immediately after the displacement in the axial direction, the trunnion 7 starts displacement in the direction opposite that in which it has been displacing, toward the neutral position. However, the trunnion 7 continues swinging around the pivot shaft 9 as long as there exists a displacement from the neutral position. As a result, a displacement of the precess cam l8 in a circumferential direction of a cam surface 21 is transmitted to the spool 15 via the link arm 19, to thus displace the spool 15 in the axial direction. Then, under a state where the inclination angle of the trunnion 7 reaches a predetermined angle corresponding to the desired transmission ratio, the control valve 12 is closed simultaneously with the trunnion 7 returning to the neutral position. Hence, supply and discharge of the pressurized oil to and from the actuators 10, 10 is stopped. As a result, the inclination angle of the trunnion 7 becomes an angle corresponding to the amount of displacement of the sleeve 14 in the axial direction displaced by the stepping motor 13.
During operation of the toroidal continuously variable transmission such that as described above, the input disc 2 (the left-hand input disc 2 in FIGS. 7 and 8) is driven and rotated by a driving shaft 22 that is connected with a power source such as an engine via a hydraulic loader 23 as shown in FIGS. 7 and 8. As a result, the pair of the input discs 2, 2 supported on the respective end portions of the input shaft 1 rotate synchronously while being pressed in a direction approaching toward each other. Then, the rotational movement is transmitted to the output discs 5, 5, via the power rollers 6, 6, and output from the output gear 4.
As described above, when the power is transmitted from the input discs 2, 2 to the output discs 5, 5, a force is applied on the trunnions 7, 7 in a direction along the pivot shafts 9, 9, which are provided on respective ends of the trunnions 7, 7, due to friction on rolling contact portions (i.e., traction portions) between the peripheral surfaces of the power rollers 6, 6 supported on the inner surfaces, and the inner surfaces of the discs 2, 5. This force is referred to as a 2Ft, and the magnitude of the force is proportional to a torque transmitted from each of the input side discs 2, 2 to each of the output side discs 5, 5 (or from the output side discs 5, 5 to the input side discs 2, 2). Such a force 2Ft is supported by the actuators 10, 10. Therefore, during operation of the toroidal continuously variable transmission, a pressure differential between a pair of oil pressure chambers 24a and 24b provided on respective sides of the pistons 16, 16 constituting the actuators 10, 10 is proportional to the magnitude of the force 2Ft.
In the case where rotational speeds of the input shaft 1 and the output gear 4 are changed, when deceleration is performed between the input shaft 1 and the output gear 4 first, the trunnions 7, 7 are moved in the axial directions of the pivot shafts 9, 9 by the actuators 10, 10, thereby swinging the trunnions 7, 7 to a position shown in FIG. 8. Then, as shown in FIG. 8, the peripheral surfaces of the power rollers 6, 6 abut against portions of the inner surfaces of the input discs 2, 2 near the center and portions of the inner surfaces 4a of the output discs 5, 5 near the outer periphery, respectively. In contrast, at the time of increasing the speed, the trunnions 7, 7 are made swung in the opposite direction to that shown in FIG. 8. Accordingly, the trunnions 7, 7 are inclined so that, in the reverse state of that shown in FIG. 8, the peripheral surfaces of the power rollers 6, 6 abut against areas located slightly toward the outer peripheries of the inner surfaces of the input discs 2, 2 and areas located slightly toward the centers of the inner surface of the output discs 5, 5, respectively. When the inclination angle of the trunnions 7, 7 is set at an intermediate angle between the above two angles, an intermediate transmission ratio (speed ratio) can be obtained between the input shaft 1 and the output gear 4.
Further, for the case where the toroidal continuously variable transmission which is constituted and functions as described above is incorporated into an actual continuously variable transmission for a vehicle, there has been previously proposed combining the transmission with a differential unit such as a planetary gear mechanism to thereby constitute a continuously variable transmission apparatus. For example, U.S. Pat. No. 6,251,039 discloses a so-called geared-neutral-type continuously variable transmission apparatus which can switch rotation of an output shaft between forward and reverse with a stop state interposed therebetween while an input shaft rotates in a single direction. FIG. 10 shows the continuously variable transmission apparatus disclosed in U.S. Pat. No. 6,251,039. The continuously variable transmission apparatus is constituted by combining a toroidal continuously variable transmission 25 and a planetary-gear-type transmission 26. The toroidal continuously variable transmission 25 is provided with an input shaft 1, a pair of input discs 2, 2, an output disc 5a, and a plurality of power rollers 6, 6. In the example shown in FIG. 10, the output discs 5a is constituted such that outer surfaces of the pair of output discs abut each other to be formed integrally.
The planetary-gear-type transmission 26 is provided with a carrier 27 which is fixedly connected on the input shaft 1 and one of the input discs 2 (the right-hand input disc in FIG. 10). A first transmission shaft 29 is rotatably supported on an intermediate portion in the radial direction of the carrier 27, and planetary gear elements 28a, 28b are fixedly disposed on respective end portions of the first transmission shaft 29. Further, a second transmission shaft 31 is rotatably supported, with the carrier 27 disposed between the second transmission shaft 31 and the input shaft 1 concentrically with the input shaft 1, and sun gears 30a, 30b fixedly disposed on respective ends of the second transmission shaft 31. Furthermore, the planetary gear elements 28a is meshed with a sun gear 33 which is fixedly disposed on a tip portion of a hollow rotary shaft 32 whose base portion (the left end portion in FIG. 10) is connected to the output disc 5a, and/or the planetary gear element 28b is meshed with the sun gear 30a which is fixedly disposed on one end portion (the left end portion in FIG. 10) of the second transmission shaft 31. The planetary gear element 28a (the left-hand element in FIG. 10) is also meshed with a ring gear 35 which is rotatably provided around the carrier 27 via another planetary gear element 34.
Meanwhile, planetary gear elements 37a, 37b are rotatably supported on a second carrier 36 which is provided around the sun gear 30b which is fixedly disposed on the other end portion (the right end portion in FIG. 10) of the second transmission 31. The second carrier 36 is fixedly disposed on a base end portion (the left end portion in FIG. 10) of an output shaft 38 which is provided concentrically with the input shaft 1 and the second transmission shaft 31. The planetary gear elements 37a, 37b are meshed with each other. Further, the planetary gear element 37a, one of the planetary gear elements, is meshed with the sun gear 30b. The other planetary gear element 37b is meshed with a second ring gear 39 which is rotatably provided around the second carrier 36. The ring gear 35 and the second carrier 36 are allowed to engage and disengage by way of a low-speed clutch 40. The second ring gear 39 and a stationary portion such as a housing are allowed to engage and disengage by way of a high-speed clutch 41.
In the case of the continuously variable transmission apparatus shown in FIG. 10 described above, under a so-called low-speed mode where the high-speed clutch 41 is disengaged simultaneously with engagement of the low-speed clutch 40, the power of the input shaft 1 in transmitted to the output shaft 38 via the ring gear 35. By changing the transmission ratio of the toroidal continuously variable transmission 25, the overall transmission ratio of the continuously variable transmission apparatus; that is, the transmission ratio between the input shaft 1 and the output shaft 38, is changed. Under such a low-speed mode, the overall transmission ratio of the continuously variable transmission apparatus changes infinitely. In other words, by adjusting the transmission ratio of the toroidal continuously variable transmission 25, a rotation state of the output shaft can be switched between forward and reverse with a stop state interposed therebetween while the input shaft rotates in a single direction.
During acceleration or during constant-speed driving under such a low speed mode, a torque passing through the toroidal continuously variable transmission 25 (hereinafter referred to as “passing torque”) is applied on the output disc 5a from the input shaft 1 via the carrier 27, the first transmission shaft 29, the sun gear 33, and the hollow rotary shaft 32. Further, the torque is applied on the input discs 2, 2 from the output disc 5a via the power rollers 6, 6. In other words, the torque passing through the toroidal continuously variable transmission 25 is circulated in a direction where the input discs 2, 2 receive torque from the power rollers 6, 6 during acceleration or constant-speed driving.
In contrast to the above, under a so-called high-speed mode where the low-speed clutch 40 is disengaged and the high-speed clutch 41 is engaged, the power of the input shaft 1 is transmitted to the output shaft 38 via the first and second transmission shafts 29, 31. By changing the transmission ratio of the toroidal continuously variable transmission 25, the overall transmission ratio of the continuously variable transmission apparatus is changed. In the above case, the higher the transmission ratio of the toroidal continuously variable transmission 25, the higher the overall transmission ratio of the continuously variable transmission apparatus.
Note that the torque passing through the toroidal continuously variable transmission 25—during acceleration or constant—speed driving under such a high speed mode—is applied in a direction where the input discs 2, 2 add torque on the power rollers 6, 6.
For example, in the case of a continuously variable transmission apparatus having such a structure as shown in FIG. 10 and capable of implementing a so-called infinitely variable transmission ratio where the output shaft 38 is stopped while the input shaft 1 rotates, it is important to maintain a torque applied on the toroidal continuously variable transmission 25 at an appropriate value under a state where the output shaft 38 is stopped and the transmission ratio is drastically increased, in view of ensuring durability and easy operability of the toroidal continuously variable transmission 25. The reason for the above is as follows. As is clear from a relation of “rotational driving power=rotation speed×torque,” under a state where the transmission ratio is extremely high and the output shaft 38 stops, or rotates at a very low speed, with the input shaft 1 rotating, the torque passing through the toroidal continuously variable transmission 25 (passing torque) becomes larger than the torque applied on the input shaft 1. Therefore, in order to secure durability of the toroidal continuously variable transmission 25 without upsizing the toroidal continuously variable transmission 25, there must be adopted strict control for confining the torque within a range of appropriate values. More specifically, a control inclusive of a driving source is required for stopping the output shaft 38 while minimizing a torque input onto the input shaft 1.
Meanwhile, under a state where the transmission ratio is extremely high, the torque applied on the output shaft 38 changes to a large extent even when the transmission ratio of the toroidal continuously variable transmission 25 is changed slightly. Accordingly, unless the transmission ratio of the toroidal continuously variable transmission 25 is adjusted strictly, the driver may experience an uncomfortable feeling, or drivability may be poor. For example, in the case of an automatic transmission for a vehicle, a stopping state is sometimes maintained while the driver steps on the brake during a vehicle stop. Under such a state, when the transmission ratio of the toroidal continuously variable transmission 25 is not adjusted strictly and a large torque is applied on the output shaft 38, a force required for stepping on the brake pedal during the vehicle stop becomes larger, thereby increasing driver fatigue. Meanwhile, when the transmission ratio of the toroidal continuously variable transmission 25 is not adjusted strictly and too small a torque is applied on the output shaft 38, vehicle start may fail to be smooth, or the vehicle may roll in reverse while starting on an uphill grade. Therefore, during a vehicle stop or very low speed driving, strict adjustment of the transmission ratio of the toroidal continuously variable transmission 25 is required in addition to control of the torque transmitted from the driving source to the input shaft 1.
In consideration of the above, JP-A-10-103461 discloses a structure where the torque passing through the toroidal continuously variable transmission (passing torque) is regulated directly through control of a pressure differential between hydraulic actuator which are used for displacing trunnions.
However, in the case of the structure disclosed in JP-A-10-103461, because the control relies on only a pressure differential, stopping a posture of the trunnion at the moment when the passing torque has reached the desired value is difficult. More specifically, because a displacement amount of the trunnion becomes large in order to control the torque, there easily occurs so-called overshoot (and hunting resulting from the overshoot). . . where the trunnion continues displacement without stopping at the moment when the passing torque becomes coincident with the target value. Hence, control of the passing torque is not stable.
In particular, the overshoot is easily introduced in a case of a toroidal continuously variable transmission 25 of a so-called “without cast angle type,” as in the case of general half-toroidal continuously variable transmissions shown in FIGS. 7 through 9, wherein a direction along the pivot shafts 9, 9 which are provided on both end portions of the trunnions 7, 7 and a direction of center shafts of the input and output discs 2, 5 are perpendicular to each other. In contrast, in a case of a continuously variable transmission whose structure includes a cast angle as in the case of a general full-toroidal continuously variable transmission, a force in a direction of converging an overshoot acts thereon. Therefore, sufficient torque control is conceivably performed even with the structure disclosed in the above-cited JP-A-10-103461.
In view of the above circumstances, there can be conceived a control method or a control device with which a torque passing through the toroidal continuously variable transmission (passing torque) can be controlled strictly even in the case of a continuously variable transmission apparatus including a toroidal continuously variable transmission without cast angle, as in the case of a general half-toroidal continuously variable transmission.
FIG. 11 shows an example structure of a continuously variable transmission apparatus having such a control method and a control device as described above. The continuously variable transmission apparatus shown in FIG. 11 has a function similar to that of a conventionally known continuously variable transmission apparatus shown in the above-cited FIG. 10; however, assembly of the planetary-gear-type transmission 26a portion is improved by contriving a structure of the planetary-gear-type transmission 26a portion.
First and second planetary gears 42, 43, both being of double pinion type, are supported on respective sides of a carrier 27a which is rotated together with the input shaft 1 and the pair of input discs 2, 2. That is, the first planetary gear 42 is constituted of a pair of planetary gear elements 44a, 44b, and the second planetary gear 43 is constituted of a pair of planetary gear elements 45a, 45b. The planetary gear elements 44a, 44b are meshed with each other, as are the planetary gear elements 45a, 45b. Further, the planetary gears elements 44a, 45a on the inner periphery are meshed with first and second sun gears 47, 48, respectively, which are fixedly disposed on a tip portion (the right end portion in FIG. 11) of a hollow rotary shaft 32a whose base portion (the left end portion in FIG. 11) is connected to the output disc 5a, and on one end portion (the left end portion in FIG. 11) of a transmission shaft 46. The planetary gear elements 44b, 45b on the outer periphery are meshed with a ring gear 49.
Meanwhile, planetary gear elements 51a, 51b are rotatably supported on a second carrier 36a which is provided around a third sun gear 50—which is fixedly disposed on the other end portion (the right end portion in FIG. 11) of the transmission shaft 46. The second carrier 36a is fixedly disposed on the base end portion (the left end portion in FIG. 11) of an output shaft 38a, which is concentrically provided with the input shaft 1. The planetary gear elements 51a, 51b are meshed with each other. Further, the planetary gear element 51a on the inner periphery side is meshed with the third sun gear 50, and the planetary gear element 51b on the outer periphery side is meshed with a second ring gear 39a, which is rotatably provided around the second carrier 36a. The ring gear 49 and the second carrier 36a are allowed to engage and disengage by way of a low-speed clutch 40a. The second ring gear 39a and a stationary portion such as a housing are allowed to engage and disengage by way of a high-speed clutch 41a. 
In the case of an improved continuously variable transmission apparatus structured as described above, under a state where the high-speed clutch 41a is disengaged simultaneously with engagement of the low-speed clutch 40a, the power of the input shaft 1 is transmitted to the output shaft 38a via the ring gear 49. By changing the transmission ratio of the toroidal continuously variable transmission 25, an overall speed ratio eCVT of the continuously variable transmission apparatus; that is, a speed ratio between the input shaft 1 and the output shaft 38a, is changed. In the above case, a relationship between the speed ratio (i.e., the transmission ratio) eCVU of the toroidal continuously variable transmission 25 and the overall speed ratio eCVT of the continuously variable transmission apparatus, where a ratio between a number of teeth m49 of the ring gear 49 and a number of teeth m47 of the first sun gear 47 is set at i1 (=m49/m47), can be represented by the following Equation 1.eCVT=(eCVU+i1−1)/i1  (1)
In the case where, for example, the ratio i1, the numbers of teeth, is 2, the relationship between the two speed ratios of eCVU and eCVT changes as shown by a line α in FIG. 12.
Meanwhile, under a state where the low-speed clutch 40a is disengaged and the high-speed clutch 41a is engaged, the power of the input shaft 1 is transmitted to the output shaft 38a via the first planetary gear 42, the ring gear 49, the second planetary gear 43, the transmission shaft 46, the planetary gear elements 51a, 51b, and the second carrier 36a. By changing the speed ratio eCVU of the toroidal continuously variable transmission 25, the overall speed ratio eCVT of the continuously variable transmission apparatus is changed. A relationship in the above case between the speed ratio eCVU of the toroidal continuously variable transmission 25 and the overall speed ratio eCVT of the continuously variable transmission apparatus can be represented by the following Equation 2. In Equation 2, i1 represents a ratio (=m49/m47) between the number of teeth m49 of the ring gear 49 and the number of teeth m47 of the first sun gear 47, i2 represents a ratio (=m49/m48) between the number of teeth m49 of the ring gear 49 and a number of teeth m48 of the second sun gear 48, and i3 represents a ratio between a number of teeth m39 of the second ring gear 39a and a number of teeth m50 of the third sun gear 50 (=m39/m50).eCVT={l/(1−i3)}·{l+(i2/i1)(eCVU−1)}  (2)
In the case where the ratio i1 is 2, i2 is 2.2, and i3 is 2.8, the relationship between the two speed ratios of eCVU and eCVT changes as shown by a line β in FIG. 12.
As is clear from the line α in FIG. 12, a continuously variable transmission apparatus which is constituted and functions in the aforesaid manner can realize a so-called infinitely variable transmission ratio state where the output shaft 38a is stopped while the input shaft 1 rotates. However, as mentioned previously, under such a state where the output shaft 38 is stopped or driven at a very low speed while the input shaft 1 rotates, the torque passing through the toroidal continuously variable transmission 25 (i.e., passing torque) becomes greater than a torque applied on the input shaft 1 from the engine which is the driving source. For this reason, the torque which is input from the driving source into the input shaft 1 must be regulated properly during a vehicle stop or during very low speed driving in order to prevent the passing torque from becoming excessively large (or excessively small).
Further, during the very low speed driving, under a state where the output shaft 38a is almost stopped; in other words, under a state where the transmission ratio of the continuously variable transmission apparatus is significantly large and the rotation speed of the output shaft 38a is significantly slower than that of the input shaft 1, the torque applied on the output shaft 38a fluctuates to a large extent upon a slight fluctuation in the transmission ratio of the continuously variable transmission apparatus. Therefore, the torque which is input also from the driving source to the input shaft 1 must be regulated properly in order to secure smooth drivability.
During acceleration or constant-speed driving under such a low-speed mode, the torque is, as is the case with the conventional structure shown in the aforementioned FIG. 10, applied on the output disc 5a from the input shaft 1 via the carrier 27a, the first planetary gear 42, the first sun gear 47, and the hollow rotary shaft 32a. Further, the torque is applied on the input discs 2, 2 from the output disc 5a via the power rollers 6, 6 (see FIG. 10). In other words, the passing torque is circulated in a direction where the input discs 2, 2 receive torque from the power rollers 6, 6 during acceleration or constant-speed driving.
For this reason, as shown in FIG. 13, a control method and a control apparatus according to the above constitution are arranged such that the torque input from the driving source into the input shaft 1 is regulated properly. First, a rotation speed of the engine which serves as a driving source is controlled roughly. Specifically, the rotation speed of the engine is regulated to a point “a” in the range of “w” of FIG. 13. In conjunction with the above, there is set the transmission ratio of the toroidal continuously variable transmission 25 which is required for matching a rotation speed of the input shaft 1 of the continuously variable transmission apparatus with the controlled rotation speed of the engine. This setting is to be operated according to the above-mentioned Equation 1. That is, the torque transmitted from the engine to the input shaft 1 must be strictly regulated in the case of a so-called low-speed mode where the low-speed clutch 40a is engaged and the high-speed clutch 41a is disengaged. Therefore, the transmission ratio of the toroidal continuously variable transmission 25 is to be set, according to Equation 1, such that the rotation speed of the input shaft 1 corresponds to the required rotation speed of the output shaft 38a. 
Meanwhile, a pressure differential between the oil pressure chambers 24a, 24b (see FIG. 9 and FIG. 15 described later) incorporated in the hydraulic actuators 10, 10—used for displacing the trunnions 7, 7 incorporated in the toroidal continuously variable transmission 25 in the direction along the pivot shafts 9, 9—is measured with an oil pressure sensor 52 (see FIG. 2, described later). The oil pressure is measured under a state where the rotation speed of the engine is roughly controlled (however, the rotation speed must be maintained constant) and, corresponding thereto, the transmission ratio of the toroidal continuously variable transmission 25 is set according to Equation 1 in the manner described above. Then, the torque passing through the toroidal continuously variable transmission 25 (passing torque) TCVU is calculated from the oil pressure differential obtained from the measurement.
Specifically, so long as the transmission ratio of the toroidal continuously variable transmission 25 is constant, the oil pressure differential is proportional to the torque TCVU passing through the toroidal continuously variable transmission 25. Accordingly, the torque TCVU can be calculated from the above oil pressure differential. The reason for this is as follows. As described above, the actuators 10, 10 support a force of so-called 2Ft having a magnitude proportional to the torque (i.e., the torque TCVU passing through the toroidal continuously variable transmission 25) transmitted from the input discs 2, 2 to the output disc 5a (or from the output discs 5a to the input discs 2, 2).
Meanwhile, the torque TCVU can be obtained from Equation 3 below.TCVU=eCVU·TIN/{eCVU+(i1−1)ηCVU}  (3)
In Equation 3, eCVU represents a speed ratio of the toroidal continuously variable transmission 25, TIN represents the torque input from the engine to the input shaft 1, i1 represents a teeth number ratio of planetary-gear-type transmission pertaining to the first planetary gear 42 (i.e., a ratio between the number of teeth m49 of the ring gear 49 and the number of teeth m47 of the first sun gear 47), and ηCVU represents efficiency of the toroidal continuously variable transmission 25.
Here, a deviation ΔT (=TCVU1−TCVU2) is obtained from TCVU1, which is the torque actually passing through the toroidal continuously variable transmission 25 as obtained from the above oil pressure differential and the target passing torque TCVU2 obtained from Equation 3. Then, the speed ratio of the toroidal continuously variable transmission 25 is adjusted in a direction where the deviation ΔT is eliminated (i.e., where ΔT becomes zero). Note that because the deviation of the torque ΔT and a deviation of the oil pressure differential are in a proportional relationship, the adjustment of the transmission ratio can be performed either by the deviation of the torque or by the deviation of the oil pressure differential. In other words, from the technical point of view, control of the transmission ratio based on the deviation of the torque is identical with control of the transmission ratio based on the deviation of the oil pressure differential.
As an example, the following is conceived under the assumption that, within the range where the actual torque TCVU1 (measured value) passing through the toroidal continuously variable transmission 25 is restricted to the target value TCVU2 as shown in FIG. 13, a torque TIN with which the engine drives the input shaft 1 changes in such a sharply decreasing direction that the rotation speed of the input shaft 1 is increased. Such characteristics of the engine can be easily obtained even in a low-speed rotation range when the engine is electronically controlled. In a case where the engine has such characteristics and where the measured torque value TCVU1 has a deviation from the target torque value TCVU2 in the direction in which the input discs 2, 2 receive torque from the power rollers 6, 6 (see FIGS. 8 through 10), the overall transmission ratio of the continuously variable transmission apparatus is displaced to the deceleration side so as to increase the rotation speed of the engine to thereby reduce the torque TIN which drives the input shaft 1. To achieve the above, the transmission ratio of the toroidal continuously variable transmission 25 is changed to the acceleration side.
However, under a vehicle stop state where the driver steps on a brake pedal (i.e., a state where the rotation speed of the output shaft is zero), the transmission ratio of the toroidal continuously variable transmission 25 is controlled within a range where the adjusted force can be absorbed by a slip generated in the toroidal continuously variable transmission 25; that is, a slip (creep) generated on the contact portions (i.e., traction portion) of the inner surfaces of the input and output discs 2, 5a and the peripheral surfaces of the power rollers 6, 6 (see FIGS. 8 through 10). Therefore, an allowable range for adjusting the speed ratio is limited to a range where strain is not applied on the contact portions, which is a stricter limitation than that imposed in the case of low-speed driving.
For example, when the target value TCVU2 is at point “a” and the measured value TCVU1 is at point “b” in FIG. 13, the input discs 2, 2 have deviation in a direction receiving a torque from the power rollers 6, 6. Here, the speed ratio eCVU of the toroidal continuously variable transmission 25 is changed to the acceleration side so that the overall speed ratio eCVT of the continuously variable transmission apparatus (T/M) is changed to the deceleration side. A rotation speed of the engine is increased in conjunction with the above so as to lower the torque. In contrast, when the measured value TCVU1 is at point “c” in FIG. 13, the input discs 2, 2 have deviation in a direction where torque is added on the power rollers 6, 6. In the case where TCVU1 is at point “c,” reverse to the case where TCVU1 is at the point “b,” the speed ratio eCVU of the toroidal continuously variable transmission 25 is changed to the deceleration side so that the overall speed ratio eCVT of the continuously variable transmission apparatus (T/M) is changed to the acceleration side. The rotation speed of the engine is decreased in conjunction with the above so as to increase the torque.
The above-mentioned operations are repeated until the torque TCVU1 actually passing through the toroidal continuously variable transmission 25 as obtained from the oil pressure differential matches the target value. In other words, the above-mentioned operations are repeated in the case where the torque TCVU1 passing through the toroidal continuously variable transmission 25 cannot be matched with the target value TCU2 through only one iteration of transmission gear control of the toroidal continuously variable transmission 25. As a result, the torque TIN with which the engine rotates and drives the input shaft 1 can be set closer to a value which allows the torque TCVU passing through the toroidal continuously variable transmission 25 to reach the target value TCVU2.
Note that the above operations are performed automatically and in a short period of time through instructions from a microcomputer which is incorporated in a controller of the continuously variable transmission apparatus.
FIG. 14 shows relationships among a ratio (the left-handed vertical axis) of the torque TCVU passing through the toroidal continuously variable transmission 25 and the torque TIN with which the engine rotates and drives the input shaft 1, an overall speed ratio eCVT (horizontal axis) of the continuously variable transmission apparatus, and a speed ratio eCVU (the right-handed vertical axis) of the toroidal continuously variable transmission 25. The solid line “a” shows a relationship between the ratio of the passing torque TCVY to the driving torque TIN and the overall speed ratio eCVT of the continuously variable transmission apparatus, and the dotted line “b” shows a relationship between the two speed ratios eCVT and eCVU. The above constitution regulates the speed ratio eCVU of the toroidal continuously variable transmission 25 so as to regulate the torque TCVU1 actually passing through the toroidal continuously variable transmission 25 to the target value (TCVU2) represented by points on the solid line “a” under a state where the overall speed ratio eCVT of the continuously variable transmission apparatus is regulated to a predetermined value.
In the above constitution, control for regulating the torque TCVU1 actually passing through the toroidal continuously variable transmission 25 to the point on the solid line “a,” which is the target value TCVU2, is performed in two stages. Specifically, the rotation speed of the engine is roughly controlled to a specific rotation speed; that is, to a value assumed to provide the target value TCVU2. Thereafter, the transmission ratio of the toroidal continuously variable transmission 25 is controlled in conjunction with the rotation speed thereof. For this reason, in contrast to the case of a conventional method, the torque TCVU1 actually passing through the toroidal continuously variable transmission 25 can be regulated to the target value TCVU2 without introducing an overshoot (and resultant hunting), or even when introduced, the overshoot is suppressed within such a level that would not raise any problems in practical use.
Note that, as described above, under a state of a vehicle stop with the driver stepping on a brake pedal, a driving force (torque) is applied on the output shaft 38a (FIG. 10) based on the slip generated within the toroidal continuously variable transmission 25. The magnitude of the torque may be set to a value which corresponds to a creep torque generated in a general automatic transmission provided with a torque converter. The reason for this is to avoid providing an uncomfortable feeling to a driver who is accustomed to operations of a general automatic transmission. In addition, a direction of the torque is determined by a position of a control lever provided at a driver's seat. When a forward direction (D range) is selected by the control lever, a torque of a forward direction is applied on the output shaft 38a. When a reverse (R range) is selected, a torque of reverse direction is applied.
Next, a circuit in a section which controls the speed ratio of the toroidal continuously variable transmission 25 so that the torque TCVU1 actually passing through the toroidal continuously variable transmission 25 matches the target value TCVU2 will be described with reference to FIG. 15. By way of the control valve 12a, pressurized oil can be supplied to and discharged from the pair of oil pressure chambers 24a, 24b included in the hydraulic actuators 10, 10 which are used for displacing the trunnions 7, 7 in the axial direction (in the vertical direction in FIG. 15) of pivot shafts 9, 9 (see FIG. 9). The sleeve 14 constituting the control valve 12 is allowed to displace in the axial direction by a stepping motor 13 via a link arm 54 and a rod 53. The spool 15 constituting the control valve 12 is engaged with the trunnion 7 via the link arm 19, the precess cam 18, and the rod 17. The spool 15 is allowed to displace in the axial direction in conjunction with a displacement in the axial direction and a swing of the trunnion 7. The above constitution is principally the same as that of a conventionally known toroidal continuously variable transmission.
The above constitution is particularly arranged such that the sleeve 14, driven by the stepping motor 13, can also be driven by a hydraulic differential pressure cylinder 55. Specifically, a tip portion of the rod 53 whose base end portion is connected to the sleeve 14 is pivotally supported by an intermediate portion of the link arm 54. Further, pins provided at output portions of the stepping motor 13 or the differential pressure cylinder 55 are engaged with elongated holes provided on respective end portions of the link arm 54. When one pin in the elongated hole provided on one of the two end portions of the link arm 54 is pushed or pulled, the other pin in the elongated hole on the other end portion serves as a pivot. According to such a constitution, the sleeve 14 can be displaced in the axial direction not only by the stepping motor 13 but also by the differential pressure cylinder 55. The above constitution is arranged such that the speed ratio eCVU of the toroidal continuously variable transmission 25 can be adjusted by a displacement of the sleeve 14 caused by the differential pressure cylinder 55 depending on the torque TCVU passing through the toroidal continuously variable transmission 25.
In order to achieve the above, the constitution is arranged such that different oil pressures can be induced into a pair of oil pressure chambers 56a, 56b provided in the differential pressure cylinder 55 via a correcting control valve 57. Oil pressures introduced into the oil pressure chambers 56a, 56b are determined from a pressure differential ΔP between oil pressures PDOWN and PUP which act in the pair of oil pressure chambers 24a, 24b constituting the actuator 10; and a pressure differential ΔPO between output pressures of a pair of solenoid valves 58a and 58b used for adjusting opening of the correcting control valve 57. Specifically, opening and closing of the two solenoid valves 58a, 58b are calculated by an unillustrated control device (hereinafter referred to as “controller”), and controlled on the basis of an output signal output from the controller such that the pressure differential ΔPO between the output pressures of the two solenoid valves 58a and 58b reaches a target pressure differential corresponding to the target torque TCVU2 of the toroidal continuously variable transmission 25. Accordingly, the following forces act on a spool 59 constituting the correcting control valve 57: a force corresponding to the pressure differential ΔP between oil pressures acting on the oil pressure chambers 24a, 24b of the actuator 10; and a pressure differential ΔPO—between the output pressures of the solenoid valves 58a, 58b—which is the target pressure differential corresponding to the target torque TCVU2; that is, counterforce against ΔP.
In the case where the torque TCVU1 actually passing through the toroidal continuously variable transmission 25 is identical with the target torque TCVU2; that is, in the case where a difference ΔT between the passing torque TCVU1 and the target torque TCVU2 is zero, the force corresponding to the pressure differential ΔP between oil pressures acting on the oil pressure chambers 24a, 24b of the actuator 10 and the force corresponding to the pressure differential ΔPO between the output pressures of the solenoid valves 58a, 58b are balanced. For this reason, the spool 59 constituting the correcting control valve 57 is brought into a neutral position, and the pressures acting on the oil pressure chambers 56a, 56b of the differential pressure cylinder 55 become equal to each other. Under the above state, a spool 60 of the differential pressure cylinder 55 is brought into a neutral position, and the speed ratio of the toroidal continuously variable transmission 25 remains unchanged (not corrected).
Meanwhile, when a difference arises between the torque TCVU1 actually passing through the toroidal continuously variable transmission 25 and the target torque TCVU2, balance is lost between the force corresponding to the pressure differential ΔP between the oil pressures acting on the oil pressure chambers 24a, 24b of the actuator 10 and the force corresponding to the pressure differential ΔPO between the output pressures of the solenoid valves 58a, 58b. Then, according to a magnitude and direction of the difference ΔT between the passing torque TCVU1 and the target torque TCVU2, the spool 59 constituting the correcting control valve 57 is displaced in the axial direction, to thus induce an oil pressure corresponding to the magnitude and direction of ΔT into the oil pressure chambers 56a, 56b of the differential pressure cylinder 55. Then, the spool 60 of the differential pressure cylinder 55 is displaced in the axial direction, whereby the sleeve 14 constituting the control valve 12 is displaced in the axial direction. Consequently, the trunnions 7, 7 are displaced in the direction along the pivot shafts 9, 9 to thus change (correct) the speed ratio of the toroidal continuously variable transmission 25. Note that the direction and the amount of the displacement of the speed ratio in relation to the above is the same as described with reference to the aforementioned FIGS. 13 and 14. The amount of displacement of the speed ratio; that is, the amount to be corrected (i.e., amount to be corrected in relation to the speed ratio) of the toroidal continuously variable transmission 25 in relation to the above is sufficiently small as compared with a speed ratio width of the toroidal continuously variable transmission 25. For this reason, a stroke of the spool 60 of the differential pressure cylinder 55 is designed so as to be sufficiently smaller than a stroke of an output portion of the stepping motor 13.
In the case where the conventional continuously variable transmission apparatus shown in FIG. 10 or the structure shown in FIG. 11 is employed as an actual automatic transformer of a vehicle, when a non-travel range is selected with a shift lever provided at a driver's seat, each of low-speed clutches 40, 40a and high-speed clutches 41a, 41b are arranged to be disengaged. Specifically, in the case where the shift lever is in a neutral range (N range) or in a parking range (P range)—each used for selecting a state where a vehicle is not allowed to drive—each of the clutches 40, 40a, 41a, and 41b is disengaged. As a result, the torque passing through the toroidal continuously variable transmission 25 and the planetary-gear-type transmission 26, 26a becomes quite small (substantially zero). Accordingly, durability of the toroidal continuously variable transmission 25 and the planetary-gear-type transmission 26, 26a can be secured.
However, in such a state where each of the clutches 40, 40a, 41a, 41b is disengaged and the torque passing through the toroidal continuously variable transmission 25 becomes quite small, the correction of the speed ratio of the toroidal continuously variable transmission 25 according to FIGS. 13 and 14 cannot be performed accurately. More specifically, the speed ratio may be corrected excessively because of a failure to control the transmission ratio (speed ratio) based on the torque passing through the toroidal continuously variable transmission 25. Further, under such a state that the speed ratio is corrected excessively, when a travel range (drive range or a reverse range) is selected with a shift lever, an excessive torque may be applied on the output shaft 38, 38a at a moment when the low-speed clutch 40, 40a is engaged. When an excessive torque is applied on the output shaft 38, 38a in this manner, a driver may feel undesirably uncomfortable feeling at a vehicle start.